Blade structure for turbines



Oct; 28, 1969 WJ ICKUHR 3,475,108

BLADE STRUCTURE FOR TURBINES v I Filed Feb. 14, 1968 (5 Sheets-Sheet 1 IW. 'ZICKUHR BLADE STRUCTURE FOR TURBINES Oct. 28, 1969 Filed Feb. 14,1968 I l l 3 Sheets-Sheet 2 Fig.5a

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| l i v I 1 1 l 6 I i Fig.5c

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Oct. 28, 1969 w. ZICKUHR 3,475,108

BLADE STRUCTURE FOR TURBINES Filed Feb. 14, 1968 3 SheetsSheet :5

[ 1 kWh v (25% m v ado AD [mm] 0 160 zbn Fig.7 "sm .u 11 Q 11 11 1 Pr3Pr Fig.8

United States Patent US. Cl. 415194 14 Claims ABSTRACT OF THE DISCLOSUREA blade structure for turbines such as high-pressure axial turbines. Theturbine has at least One ring of stationary blades and a rotor whichcarries a coac'ting ring of rotary blades, these blades forming at leastone of a series of turbine stages connected one next to the other in thedirection in which the pressure of the turbine-driving pressure fluiddrops, while this fluid expands during driving of the rotor. The bladeshave at a mean diameter at a turbine region where a given operatingfactor prevails and at the rated capacity of the turbine stage, with aReynolds number of at least 2x10 an opening value e=e/ t which is atleast equal to 0.5 as an average value for the blades where e is thesmallest distance between adjoining blades and t is the distance betweencorresponding points of a pair of adjoining blades. This operatingfactor at the latter turbine region is either a pressure which is equalto or greater than 2 atmospheres absolute or a region of the turbinewhere there is an increase of not more than in the specific volume ofeach stage.

My invention relates to turbines and in particular to blade structurefor turbines.

Turbines such as steam turbines are known in the most widely differentforms. In their basic construction they include at least a stationaryring of blades and a rotor which carries a ring of blades for rotationwith respect to the stationary ring of blades. The adjoining stationaryand rotary rings of blades form generally at least one of a series ofturbine stages located one after the other in the direction in which thepressure of the steam or other pressure medium drops, while thispressure fluid expands to drive the rotor. 1n the design anddimensioning of such turbines it is always attempted to make use of thestrength of the materials in conjunction with the greatest possibleoperating safety while maintaining construction costs as low as possibleand achieving at the same time a good efficiency.

It is an object of my invention also to achieve these objectives, whichas to say to provide a turbine, particularly a steam turbine, which iscontrasted with conventional steam turbines will have an improvedefficiency of operation and at the same time a reduction in theconstruction costs.

My invention is based upon recognition of the fact that these objectivescan be achieved by a proper selection of the primary turbine dimensions,such as, for example, the shaft diameter, the critical rotary speed at agiven operating speed as well as by the proper selection of the openingvalue e=e/ t for the blades. In this latter equation t represents thecircumferential distance between corresponding points of a pair ofadjoining blades, which is to say the blade distribution of a given ringof blades, while e represents the smallest distance between a pair ofadjoining blades at the mean entrance diameter of the operating fluid.The term mean entrance diameter is explained further below.

In particular my invention is based upon recognition of the fact thatthe above-mentioned opening value of the blades, is of particularsignificance with respect to the best possible operating efficiency andconstruction cost of the turbine. Thus, it is a primary object of myinvention to ice construct the blades in such a way and to provide themwith such an opening value that the result will be an improvement in theoperating efficiency together with a reduction in the constructioncosts.

Up to the present time it has been customary to provide at the meanblade diameter an opening value 6 which at a maximum is 0.4 when aparticular turbine stage is designed at its rated capacity to operate atpressures equal to or greater than 2 atmospheres absolute. It is to beunderstood that the mean blade diameter, or admission diameter,signifies the mean diameter of that turbine stage where the pressuredrop is equal to the mean value of the pressure drops of the severalstages of a given turbine casing. The mean admission diameter isapproximately equal to the inner stream dimension plus the blade height.Thus, Traupel states in his book Thermische Turbomachinen (1st volume,Springer Publishing Co., 1958) at page 174, that for a high pressureturbine where the degree of reaction r=0.5, the best value for the anglea (entrance angle of the absolute entrance speed 0 of the steam, whichat a degree of reaction of 0.5 equals the exit angle 5 of the relativesteam exit speed W2) should be 17". On page 330 of the same book,Traupel points out that at a high pressure region of a steam turbinesmall volumes are accompanied by small angles. Appropriate design datafor the optimum points of blades follow for a conventional constantpressure stage for an entrance angle a =14 and for an exit angle 5 :20";for a conventional high pressure stage a =fi =17 30', which correspondsto an opening value e=e/t of approximately 0.3, since e=e/t isapproximately equal to the sine of the angle 8 These recommendations foran extremely small opening value have clearly their basis in the factthat it is assumed that an increase in the opening value must result inan unacceptable decrease of the circumferential operating efficiency andbecause it is assumed that with large opening values the blades are veryshort and as a result the necessarily accompanying increase in the bladetip loss 1' (flow loss at the blade tips) introduces a furtherefficiency reducing factor.

In another known text on turbine construction (Fliigel, DieDampfturbinen 1931, page 190) it is recommended that normally with drumstages fgzx =fgfl between 0.25 and 0.40 should be selected. Furthermore,Fliigel writes that with sharp volume increases during progressive eX-pansion, particularly in low pressure regions it is essential to resortto steeper exit inclinations so that there will be no excessively rapidincrease in blade length which would be undesirable for the flow, andthus at high pressure stages tga =O.60 to 0.70 is used.

In another book on turbine construction (Bauer: Der Schiffsmaschinebau,1927, pages 65-67 as well as and 151) there is given only for the laststage of condensing turbines tgfl-values between 0.45 and 1.5, whereopening values of approximately 0.40 to 0.72 are present.

According to my invention it has been recognized that an increase of theopening value e/t above 0.4 is desirable not only in the last stages oflow pressure regions with large volume increases of approximately 200%and more at each stage, but also an improvement in the operatingefliciency and a reduction in the construction costs can be achieved ina most unexpected manner with an increased opening value for high andintermediate pressure regions of turbines where there is only a volumeincrease of approximately 3% from stage to stage at the high pressureregion of the turbine up to a volume increase of 15% from stage to stageat the intermediate pressure or mean pressure region of the turbine.

Thus, it is an object of my invention to provide blading for steamturbines and the'like where coacting rings of stationary and rotaryblades form at least one of a series of successive turbine stages in thedirection of pressure drop of the rotor driving the steam. My inventionresides in providing, at pressures at least equal to 2 atmospheresabsolute or at an increase in the specific volume from stage to stage ofless than 15%, at the rated capacity of the turbine stage, and with aReynolds number an opening value e e/l at the means blade diameter whichis equal to or greater than 0.45 as a mean value for the stationary androtary blades, where e=the smallest distance between adjoining blades,

t=the distance between corresponding points of a pair of blades,

w =the relative exit speed,

b=the blade width, and

1/*=the kinematic viscosity.

The requirement that the Reynolds number be at least equal to or greaterthan 2 10 is based upon the fact that the increase in the deflectionimpulse of a grid shovel, which is to say with a given deflection theincrease in the amount of flow throughput and thus the increase in theopening value, is comparable with the increase in the lift coefiicientof an aeroplane wing. Such an increase in the lift coefiicient is onlypossible if the drag of the primary flow stream at the boundary layer,which is to say the Reynolds number, is sufiiciently great. In thisevent there will be the required impulse exchange between the primaryflow and the boundary layer, and separation or breaking down of theboundary layer cannot take place. Developments carried out in connectionwith my invention have shown, in harmony with these considerations, thata blade profile having a large relative resistance moment, which islarge enough for the stage load, and a large opening value can only beachieved above a predetermined minimum Reynolds number, nomely 2 10 atthe rated capacity.

My invention has recognized the fact that as a result of the features ofmy invention the height of the blades, both for the stationary androtary blades, can be reduced. Inasmuch as the centrifugal forcestresses at the root of the blades for the rotary blades is reduced inthis way, the possible bending stresses for a given rotary bladeincreases, which is to say, however, that the blade width as seen in theaxial direction and thus the axial dimensions of the entire blading canbe reduced. From this factor it follows that inspite of the requiredincrease in the number of stages and cost of a unit of axial length forthe purpose of maintaining constant the critical speed, there isnevertheless a decrease in the construction cost on the order of 520% ascompared to conventional turbines and blading. Together with thisdecrease in the construction costs there is an appreciable increase inefiiciency which is on the order of l*5% of the total operatingefficiency as compared to conventional turbines, which is to say atleast of the theoretically possible increase in efiiciency.

Thus, it becomes possible with the features of my invention to providesmaller blades, to reduce the construction costs of components such asshafts, casings, blades, and at the same time to achieve an increase inefficiency.

On the basis of flow tests it has been determined that the blade tiploss in a first approximation is proportional to the ratio profile widthblade length My invention has, therefore, recognized that with full useof the maximum possible axial blade length with a given shaft diameter,taking into consideration the critical speed, the blade Widthwithgeometrically similar profiles-can be reduced with an increase in theopening value 2/! in such a way that the ratio b/l=profile width toblade length can be maintained smaller and thus the blade tip loss canbe reduced.

More accurate tests, referred to in greater detail below and taking intoconsideration the gap, piston, and blade tip loss as well as theinfluence of Parsons number and the opening value on the efficiency,show that depending upon the extent of gap loss and the proportion ofcentrifugal force stressing to the entire blade stressing, the turbineefiiciency is at a maximum when the opening value, or the opening ratioas a mean value for the stationary and rotary blades, is between 0.50and 0.75. In the region of greater opening values, which are greaterthan the preerred region between 0.65 and 0.7, there is a limit wherethe increase in efficiency can no longer be achieved, but even at thesegreater opening values there is also an increase in the savings ofconstruction costs, so that the designer can decide in this area on acompromise between reducing construction costs and increasing efiiciencyin accordance with which of these advantages he considers to be 'ofgreater importance. Therefore it is possible that under certainconditions it will be desirable to exceed an opening value of 0.7.

My invention can be used advantageously for turbines having one or morecasings, whether the turbines are condensation turbines or the reactiontype of back pressure turbines. My invention makes it highly desirableto use high pressure stages as well as to be used in the high pressureregions of steam turbines where a mean pressure on the order ofatmospheres absolute prevails. My invention furthermore can be favorablyused with steam turbines where there is a mean pressure at intermediatepressure stages on the order of 8 atmospheres absolute. Withmulti-casing turbines it is possible to use the features of my inventiononly for the high pressure part, only for the intermediate pressurepart, or for both of these parts of the turbine.

A preferred construction of the blading of my invention can be used withadvantage with steam turbines where there is an extent of reaction ofbetween 25 and 65% (reaction turbines or turbines with mixed featureswhere there is a compromise between reaction and impulse turbines),without however being limited to these particular types of turbines. Areaction or high pressure turbine is for a given construction cost,especially with large outputs, superior to a constant pressure turbineby approximately 1-1.5 in efliciency.

According to a further feature of my invention the blading is used insteam turbine where the extent of reaction is approximately 50%, whichis to say between 0.45 and 0.55 (reaction or high pressure turbine inthe narrower sense). Such blading is of advantage particularly because,as shown by speed diagrams, the entrance and exit flow angles of theblading and thus the profiles for the stationary rotary blades areapproximately the same. This simplifies the manufacture and provides auniform load on the ring of blades as seen in the direction in which thepressure within the turbine drops. With steam turbines where there is areaction extent different from 0.5, it is preferred to define the meanopening value sine 01 sine a; e z

It is also possible to use the features of my invention with advantagefor this type of turbine, as mentioned above. Steam turbines suitablefor using the blading of my invention are preferably axial turbines,which is to say all turbines where stationary and rotary blade ringsfollow alternatively one after the other along the axis of the turbinerotor.

In accordance with a further feature of my invention, the blades areconstructed in such a way that the relative maximum profile thickness ofthe blades d /L is between 0.3 and 0.4, where d is the diameter of theblade profile at the inner circuit at its thickest region and L is thechord length of the profile (the longest dimension or thecross-sectional blade profile). This feature provides a large strengthfor the blades because of the high resistance moments, and in additionwithin this region there is the greatest efiiciency.

A further increase in efliciency can in particular be achieved withblading having appreciable centrifugal force stresses for the rotaryblades, especially for intermediate pressure regions of turbines, byproviding the rotary blades with a tapered construction. Such a bladeconstruction differs from a cylindrical-radial blade where the crosssection of the blade is constant over the entire blade height and thecenter of gravity is situated at a radius which extends perpendicularlyfrom the rotary axis, in that the blades taper and become of graduallysmaller cross section toward their outer tips. Such an expedient resultsin a smaller width for the ring of rotary blades and thus a decrease inthe gap, piston, peripheral, and discharge losses. An increase in theflow losses does not take place. With constant relative profilethicknesses at the blade root the tapered blades will have a reductionin flow loss at the intermediate regions of the blades, since therelative profile thickness of the blade cross section becomes graduallysmaller as the cross section becomes more distant from the blade root.With tapered blade configurations it is possible to combine withincertain limits the advantages of relatively thick and relatively thinprofiles, which is to say small blades and small flow losses.

It is even better according to a further embodiment of my invention toprovide rotary blades which are twisted as well as tapered, so that itbecomes possible in this way to better adapt the profile at theindividual blade sections to the particular flow angle, and thus selectan even smaller profile thickness. Also, the extent of taper, which isto say the reduction in centrifugal force stresses as compared tocylindrical-radial blade forms, is in this case greater than the warpedblade forms of taperedtwisted construction. The strength and flowadvantages of the tapered-non-twisted and the tapered-twisted bladeforms can also be incorporated to a large extent in the rotary blades.

Further features and advantages of my invention as well as the manner ofoperation thereof are described below in connection with severalembodiments and diagrams illustrated in the drawings which accompany andform part of this application and in which:

FIG. 1 is schematic simplified illustration of a developed portion ofblade rings with stationary or rotary blades;

FIG. 2 is a schematic illustration of the section of the blades of highpressure turbine of the axial type where the blades of one rotary ringfixed to the rotor are shown in conjunction with axially adjacent bladesof a pair of adjacent stationary rings fixedly carried by the turbinecasing;

FIG. 3, which corresponds to the illustration of FIG. 1, serves toexplain the speed diagram of FIG. 4 and to derive the opening value e=e/t, where a ring of rotary blades and a ring of stationary blades areshown adjacent each other in the axial direction;

FIG. 4 is a speed diagram in which the entrance and exit angles are alsoenlarged in accordance with the enlarged opening angle of my invention;

FIGS. 5a-5d are a series of diagrams for explaining the technicaladvances achieved with my invention, and in all the diagrams 5a5d thepertinent magnitudes have been normalized into a dimensionless formdepending upon the opening value e shown at the abscissae, thisnormalizing being achieved by referring all magnitudes to acorresponding value where e=O.3, which for magnitudes having dimensionsis indicated by a dash, and in particular FIG. 5a shows a curve of thetotal efiiciency of the turbine A17=Zv;

FIG. 5b shows the curve of individual losses, and thus the piston loss 1the gap loss 11 the blade tip loss r and the e-dependant loss;

FIG. 50 shows the increase in the permissible bending stress fl'Bi andthe corresponding or normalized stage number Z;

FIG. 5d shows the drop of the corresponding blade height h, thecorresponding active length L which is to say the part of the turbineshaft where the blading is situated, the drop of the mean blade diameter5, which is analogous to the'curve h, the drop of the correspondingblade width b as well as the ratio o'F/zr which is to say the ratiodetermined by the centrifugal force at a given stress in tension of theblade root with respect to the permissible tension stress at the bladeroot;

FIG. 6 shows, on the basis of the operating efficiency improvementaccording to FIG. 5a, the attainable improvement in heat consumption ina diagram depending upon the increase in the weight diameter in thesecond casing of the turbine, which is to say at its intermediate ormean pressure casing. In this case the extent of improvement of heatconsumption indicated at the ordinate in kcal./kwh. signifies thereduction of the required generating heat for the steam per kwh. outputof electrical power, and the weight diameter, indicated on the abscissaand shown with D MD, signifies the diameter of a shaft section of theaxial length of the first turbine stage whose weight is equal to theweight of the shaft and rotary blade ring present in the first stage.This imaginary magnitude of weight diameter has proved to be ofadvantage for turbine calculations, particularly for large turbines,where beside shaft bending stresses the tangential stresses and theinfluence of the temperature field in the shaft to determine the bestpossible flow dimensions are also of significance. The use of thediameter as a magnitude influencing the shaft stressing is also properin connection with the significance of the radial dimensions of thecasing for evaluation of operating safety (temperature stressing atcasing walls and partial gap flanges). FIG. 6 shows a pair of curvedgroups, in the one case with the critical speed n =0.8 A=const. and inthe other case with n =A=const. In both groups of curves there is withincreasing rotary speeds I' which is explained further below, and forincreasing mean opening values according to sin ot =0.3, 0.4, 0.5 and0.6 an indication of the improvement in heat consumption. In this casethe curves are shown for a high pressure turbine of the axial typehaving an extent of reaction of 50% and another with the same meanentrance angle a and means exit angle FIG. 7 is a diagram of theimprovement in heat consumption with two other variations of bladeconstruction depending upon the increase in weight diameter and thus thelower curve of FIG. 7 shows the improvement in heat consumption for ablade of cylindrical-radial profile and the upper curve the improvementin the consumption for a blade having a tapered-twisted profile; and

FIG. 8 schematically illustrate-s the corresponding profileconfigurations.

Referring now to the drawings, FIG. 1 shows the section of blades for anaxial turbine of the high pressure type provided with stationary orrotary blades L, L, where the stationary blades are to be considered asuniformly distributed about the inner peripheral surface of anunillustrated turbine easing or the rotary blades are to be consideredas uniformly distributed about the exterior periphery of a turbinerotor. The distance e indi cates, as indicated, the smallest distancebetween a pair of adjoining blades and the distance 13 indicates the distance between corresponding points of a pair of adjacent blades. Theradial boundary planes 1 and 2 show where the axially adjoining nextring of rotary or stationary blades is situated, while d is the diameterof the circle which is characteristic of the maximum relative profilethickness and L is the profile chord length (the longest dimension ofthe blade profile).

The simplified mean or meridian sectional illustration of FIG. 2 showsthe rotary blades L of a rotarv rin of blades and the adjoining axiallydisplaced stationary blade rings L in both directions for a pair ofstationary blade rings. The stationary rings of blades L are fixed to acasing which is not further illustrated or to a blade carrier 3 at theblade roots 4 of the stationary blades, while the rotary blades L arefixed at their roots or feet 5 to a rotor 6 which is not furtherillustrated. Thus, as is apparent from FIG. 2, the structure shown is ahigh pressure turbine of the axial drum type. The height or longitudinallength l of the blades is indicated in FIG. 2 as well as the profilewidth 11 thereof in the axial direction.

FIG. 3 shows, in correspondence with FIG. 1, a stationary ring of bladesL and an axially adjoining rotary ring of blades L, both rings beingshown in section only with a pair of blades. The blades have entranceedges 7 and exit edges 8. The steam enters, as is known, from the ringof stationary blades L with an absolute entrance speed c and at acorresponding entrance angle as well as with a relative entrance speed Wand the corresponding entrance angle ,6 into the rotary ring of blades Lto give up at the latter a part of its kinetic energy so as to drive therotary ring of blades and the rotor, to achieve in this Way the rotaryspeed u. As is known the steam expands in the stationary ring of bladesL from the initial pressure 2 up to the final pressure 11 which inaccordance with i, the s-diagram of enthelpy difierence i -i =Hcorresponds to and provides an isentropic or theoretical entrance speedc /2gH where g indicates the acceleration of gravity. In accordance withwhether the structure forms part of a high pressure or reaction turbineor a constant pressure or impulse turbine, the steam will expand more orless also in the rotary blades L, which is to say it expands from thepressure p to the pressure p corresponding to an enthalpy differentialof i -i =H Whether the structure forms part of a high pressure or aconstant pressure turbine, it has a characteristic for the latterfactors the extent of reaction.

, =i=ra L'+ L H s where H is the drop from stage to stage (the drop froma stationary ring of blades to the adjoining rotary ring of blades).With high pressure turbines there is conventionally a degree of reactionon the order of r=50%, for example, while with action turbines there is,for example, a reaction extent of r=%. With action turbines there isthus a reversal of pressure at steam speeds encountered primarily in thenozzle rings. Thus, the invention is not limited to the magnitude of thedegree of reaction, although it is of particular advantage when usedwith high pressure turbines having an extent of reaction on the order of50%.

Returning now to FIG. 3 after the steam has given up part of its kineticenergy to the rotary ring of blades L while being deflected in thepassages 9 between the rotary blades, it leaves the latter with anabsolute exit speed c (exit angle a and the relative exit speed W2 (exitangle 5 in order to then enter into the adjoining unillustratedstationary ring of blades, or to discharge from the turbine casing, forexample for the purpose of being conducted to an adjoining turbinecomponent or to be extracted. From FIG. 3 it is clearly apparent thate+s=t-sine 5 if s indicates the width of the exit edge 8, which is tosay where e is the opening ratio or the opening value.

In accordance with the above considerations, the speed diagram of FIG.4, which is indicative of a reaction turbine having a degree of reactionon the order of r=0.5, can be readily understood. Because of theincrease in the opening value c achieved with my invention, there isprovided the illustrated increase in the mean entrance angle 04 and theincrease in the mean exit angle [3 In the illustrated example theentrance angle a is equal to the exit angle p By way of the index 112 itis indicated that the angle is the mean value for the stationary androtary blades and further that in the case of tapered or taperedtwistedblades it is also the mean value taken over the height of the blade. Inorder to better illustrate my invention there are also shown in dottedlines in FIG. 4 the angle u fl as being approximately equal to 26corresponding to an opening value 6=0.45. The vectors of the absoluteentrance and exit speeds of the stream, the relative entrance and exitspeeds, and the rotary speed are also indicated at 0 c W1, W2, and the urespectively.

My invention has recognized the fact that a blade profile with largerelative section modulus with a large opening value according to myinvention can only be developed above a predetermined minimum Reyonoldsnumber, namely 2X10 at the rated capacity. In this case the flow remainswithin the region of turbulence where the frictional drag of the primarystream at the boundary layer is sufiiciently great.

The illustrations of FIGS. 5a-5d show a constant normalized sectionmodulus for the blade profile which is based on a dimensionless bladeprofile magnitude b which is to say these figures show tests of aprofile family of the same constant resistance moment W/b Furthermore,these curves are derived with a blade profile where the quotient of madivided by d' l o.7 with an opening value 6:0.3. Furthermore,corresponding practical factors have been taken into consideration, inthat the gaps are proportional to the mean shaft diameter D the bearingdistance L is proportional to D which is to say the root from the meanadmission diameter and thus the critical speed n is approximatelyconstant. Furthermore, it is assumed that the change of the activelength of the rotor, which is to say that length which is provided withblades, equals the change in the distances between the bearings. Finallythe Parsons number X is assumed to be constant.

As is shown in FIG. 5a, there takes place in a most surprising way withan increase in the opening angle 6 an improvement in the entireefficiency to such an extent that the'sum of all of the losses Zu htransferred to the ordinate in percentage, is smaller and has itsminimum at approximately 5:0.6. This minimum can, depending upon theselected profile family, be displaced upwardly and downwardly to a smallextent.

In FIG. 51) there are indicated the improvements in percentage of thepertinent individual losses depending upon the opening value 6. And thusthere is the piston gap loss 11 indicated by the radial gap between thecompensating pistons and easing of a predetermined amount of leakageflow. With high pressure turbines, as is known, it is necessary toprovide such compensating pistons to equalize the axial thrust.

The blade gap loss a is derived from the fact that the radial gapbetween the blades and easing will result in a certain amount of leakageflow. The blade tip loss v is determined by the flow loss at the tips ofthe blades. The entrance flow value v, is determined by the followingconsideration: if the opening angle 6 is increased, then with constantblade losses there will be a decrease in the circumferential efiiciency,as already explained. On the other hand as result of the decrease in thedeflection which then takes place in the passages between the bladesthere is a decrease in the blade loss. The influence of both of thesechanges on the efiiciency is indicated by the entrance flow value 11,.An important discovery of my invention resides in the fact that, asshown in the diagram, with increasing there is not only a decrease inthe piston gap loss 11 and the gap loss 11 but also there is a decreasein the blade tip loss v so that the reduction in the individualefiiciency v, initially has no influence on the entire efiiciency and ismore than compensated for, and in fact it is only at relatively largee-values that there is an actual reduction in the total efficiency. Itis important that the blade tip loss v be proportional to the ratio of bto l, which is to say the axial blade width to the blade height or bladelength. The behavior of these individual losses provides the curvesshown in FIG. 5a for the total efficiency.

FIG. 5a shows that it is possible to reduce the blade height 75 or 1because the increased opening value provides a larger passage sectionfor the steam between the blades. At the same time it is possible toreduce the mean admission diameter 5. For the rotary blades, based upona predetermined above-mentioned centrifugal force stressing at e=0.3, itis clear that the blade width b can be reduced, because the centrifugalforce stresses and thus the ratio CF to a becomes smaller and for thisreason, as shown in the diagram of FIG. 50, the permissible bendingstresses (1 increases at the blade roots. Of course, because of theconstant Parsons number it is necessary also to increase the numbers ofstages, as shown by the curve Z in the diagram of FIG. 5c, but theextent of reduction of blade width is greater, and it is possible toreduce the active blade length of the rotor L as may be seen from FIG.5d. Therefore, since not only the mean admission diameter but also theactive length of the rotor are capable of being reduced, there is ahighly significant saving in the cost of construction.

Referring now to FIG. 6, there is indicated, as mentioned above, theimprovement in heat consumption which can be achieved in a highpressure, intermediate pressure, and low pressure casing of a 1000MW-turbine, depending upon the weight diameter in the second casingDGIMD, indicated in mm., further depending upon the opening value ordifierent sizes of the entrance angle sine a '=SiI1e {3 with a pair ofdifferent critical speeds as parameters. The further indicated pressurenumbers P are defined by the dimensionless quotients 2 A is st/ 11 whereA1, st represents the isentropic enthalpy diiterential of a given stageand a signifies the square of the circumferential speed. The pressurenumber I is, in the same way as the Parsons number X a characteristicmagnitude of the turbine and a determining factor for the stage load.The product of X and ta is, as is known, equal to 8380 m kg/ From 6 itcan be seen that with a constant critical speed and constant weightdiameter of the first stage, which is to say with constant entrance fiowmagnitudes, the operating safety, with increasing mean discharge flowangles 8 increase the gain in heat consumption. Since the distance ofthe individual curves for sine a =sine fi =0.3, 0.4, 0.5, 0.6 withincreasing 6 becomes constantly smaller, it is clear that with e-valuesabove 0.6 a saturation has been achieved which is to say in efiiciencyhas been reached where in the illustrated example it is at a maximum atvalues of 6 above 0.6. It the weight diameter is increased or thecritical speed decreased, then, as is shown in FIG. 7, with a constantdischarge flow angle the stage loading Will become reduced and the gainin heat consumption will be increased. In order to arrive at an optimumefficiency, it is recommended in accordance with my invention that therelative maximum profile thickness of the blades d /L be between 0.3 and0.4, where ai is the diameter of the blade profile where the innercircle is situated at its thickest region and L is the profile chordlength (the longest dimension of the blade section), as indicated inFIG. 1. In this way it is possible to achieve a highly as indicated inFIGS. 7 and 8, by using, instead of cylindrical-radial blades Pr atapered, or better yet a taperedtwisted blade Pr As is shown in FIG. 8the cylindricalradial blade Pr is characterized by the fact that theblade cross section IIII and the blade cross section II is constantthroughout the entire height of the blade and the center of gravitythereof is situated on a radial line which extends perpendicularlyacross the axis of rotation. Thus, these sections of the blades shown atthe left of FIG. 8 for the blade Pr are equal to each other. A taperedblade is characterized by the fact, as shown for the blade Pu in FIG. 8that the cross section of the blade diminishes gradually toward the tipthereof, so that the cross section in the region of the tip of the bladeis smaller as indicated by the same sections taken for the right bladeof FIG. 8 as compared to those taken for the left blade of FIG. 8. Atapered-twisted blade construction provides still further advantageousfeatures in that the individual sections are turned one relative to theother as indicated by the arrow at the upper right portion of FIG. 8which shows a turning of one blade section relative to the other toprovide the twisted blade structure. As a result of this piston bladestructure, as compared to the cylindricalradial structure of FIG. 8,Where the blade has a constant cross section throughout its length, itis possible to achieve the improved heat consumption indicated in FIG. 7where with the same type of illustration as shown in FIG. 6 D MDindicates again the weight diameter in the second or intermediatepressure casing and 1 indicates the pressure number. It is seen,therefore, that the blade structure Pr achieves a highly significantimprovement in heat consumption as compared to the blade structure PrThe critical rotary speed again is maintained constant. In order toindicate the changes in the gap-piston and discharge losses in a clearermanner, the improvement of the flow losses resulting from the use oftaperedtwisted blades, which results in the greatest fraction of thetotal improvement, is not shown in FIG. 7. The actual improvement ofheat consumption is therefore greater than illustrated. A tapered oreven better a taperedtwisted blade structure Pr is thus to be used withadvantage in the structure of my invention. Even if such tapered-twistedblades for reasons of strength are preferably used with advantage forrotary blades of relatively great centrifugal force stresses,nevertheless it is preferred to use such blade structures also forrotary blades subjected to lesser extents of centrifugal force stressesas well as for stationary blades because of the improvement inefficiency resulting from the use of such blades.

My invention is not to be considered as limited to use with blades forturbines where steam is the operating fluid under pressure, since it isalso capable of being used with turbines where instead of steam theoperating fluid under pressure is, for example, ammonia (NH or Freon,for example Freon 21 (CHClF Thus, the range of uses of my invention areonly limited by such factors as for a given rated capacity of a turbinestage the increase in the specific volume of each stage from one to thenext is at a maximum 15% and the Reynolds number is equal to or greaterthan 2x10 I claim:

1. In a turbine, at least one stationary ring of blades and a rotorcarrying a rotary ring of blades for rotary movement relative to thestationary ring of blades, said stationary and rotary rings forming atleasrt one of a series of turbine stages arranged one after the other inthe direction in which the pressure of the turbine-driving fluid dropswhile the turbine-driving fluid expands during rotary driving of therotor, said blades having at their mean diameter at a region of a givenoperating factor and a Reynolds number Re=w -b/v*;2 10 at the ratedcapacity of the turbine stage, an opening value e=e/t at least equal to0.45 as the mean value for the stationary and rotary blades, where e=thesmallest distance between adjoining blades,

1 1 t=the distance between corresponding points of a pair of adjoiningblades, w =the relative exit speed, b=the blade Width, and v*=thekinematic viscosity.

2. The combination of claim 1 and wherein said given factor at saidregion is a pressure which is at least equal to 2 atmospheres absolute.

3. The combination of claim 1 and wherein said given factor at saidregion is an increase of not more than 15% in the specific volume ofeach stage.

4. The combination of claim 1 and wherein the ratio 6:6/[ is between0.50 and 0.75.

5. The combination of claim 4 and wherein said ratio is between 0.60 and0.70.

6. The combination of claim 1 and wherein a highpressure region of theturbine is designed to operate at a mean pressure on the order 90atmospheres absolute.

7. The combination of claim 1 and wherein an intermediate mean region ofthe turbine is designed to operate at a mean pressure on the order of 8atmospheres absolute.

8. The combination of claim 1 and wherein the extent of reaction of theturbine is between 25 and 65%.

9. The combination of claim 8 and wherein the extent of reaction is onthe order of 50%.

10. The combination of claim 1 and wherein the turbine is an axialturbine.

11. The combination of claim 1 and wherein the relative maximum profilethicknessc of the blades d /L is between 0.3 and 0.4, wherein d is thediameter of blade profile at the inner circle arranged at its thickestregion and L is the longest chord length of the blade profile.

12. The combination of claim 11 and wherein at least the rotary bladesare tapered.

13. The combination of claim 12 and wherein at least said rotary bladesare also twisted.

14. The combination of claim 1 and wherein a fluid under pressure whichcoact with said blades is selected from the group consisting of steam,ammonia (NH and Freon 21 (CI-IClF References Cited UNITED STATES PATENTS1,539,395 5/1925 Losel 25377 1,777,098 9/ 1930 Lysholm.

FOREIGN PATENTS 1,018,407 10/ 1952 France.

1,264,388 5/1961 France.

EVERETTE A. POWELL, 111., Primary Examiner 3 UNITED STATES PATENT OFFICECERTIFICATE OF CORRECTION Patent No. 3, l-75,1O8 Dated October 213 1969Inventor(s) Waldemar Zickuhr It is certified that error appears in theabove-identified patent and that said Letters Patent are herebycorrected as shown below:

I In the heading; to the printed specification line 3,

"Walemar" should read. Waldemar SIGNED AND SEALED JUN 9 197 (SEAL)Attest:

Edward member WILLIAM 2. mm, m. Attesting Officer Commissioner ofPatents

